Variable speed liquid refrigerant pump

ABSTRACT

The invention entails the use of a positive displacement pump (41) magnetically coupled to a drive motor (42) located in a conduit arrangement (60) which is parallel to the liquid line (22) of the refigeraton system as in FIG. 5. This parallel conduit arrangement (60) also includes a pressure regulating valve (45) that will regulate the amount of pressure added to the liquid line (22) by the parallel pump and piping arrangement (60). In addition, a check valve (47) is located in the liquid line (22) to maintain the pressure differential added to the liquid line. This parallel piping arrangement (60) is desirable in order to allow a constant predetermined pressure to be added to the liquid line regardless of variations in flow rate of the liquid refrigerant. The present invention involves the use of a controlled variable speed drive on the pump meter so the flow rate through the pump will more closely match the variable system flow.

CROSS REFERENCE TO RELATED APPLICATIONS

This application is a continuation of application, Ser. No. 08/548,690,filed Oct. 25, 1995, now U.S. Pat. No. 5,749,237, which was acontinuation-in-part of application Ser. No. 08/380,739, filed Jan. 30,1995, now abandoned, which was a continuation of application Ser. No.08/127,976, filed Sep. 28, 1993, now U.S. Pat. No. 5,435,148.

This application is also the National Stage of Patent Cooperation Treatyapplication Ser. No. PCT/US96/17147, filed Oct. 26, 1996.

STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT

This invention has created without the sponsorship or funding of anyfederally sponsored research or development program.

1. Field of the Invention

This invention generally relates to the field of mechanicalrefrigeration and air conditioning and more particularly to improvingefficiency of compression-type refrigeration and air conditioningsystems.

2. Background of the Invention

In the operation of commercial freezers, refrigerators, air conditionersand other compression-type refrigeration systems, it is desirable tomaximize refrigeration capacity while minimizing total energyconsumption. Specifically, it is necessary to operate the systems at aslow a compression ratio as possible without the loss of capacity thatnormally occurs when compressor compression ratios are reduced. This isaccomplished by suppressing the formation of "flash gas" Flash gas isthe spontaneous flashing or boiling of liquid refrigerant resulting frompressure losses in refrigeration system liquid refrigerant lines.Various techniques have been developed to eliminate flash gas. However,conventional methods for suppressing flash gas can substantially reducesystem efficiency by increasing energy consumption.

FIG. 1 represents a conventional mechanical refrigeration system of thetype typically used in a supermarket freezer. Specifically, compressor10 compresses refrigerant vapor and discharges it through line 20 intocondenser 11. Condenser 11 condenses the refrigerant vapors to theliquid state by removing heat aided by circulating fan 31. The liquidrefrigerant next flows through line 21 into receiver 12. From receiver12, the liquid refrigerant flows through line 22 to counter-flow heatexchanger (not shown). After passing through exchanger, the refrigerantflows via line 23 through thermostatic expansion valve 14. Valve 14expands the liquid refrigerant to a lower pressure liquid which flowsinto and through evaporator 15 where it evaporates back into a vapor,absorbing heat. Valve 14 is connected to bulb 16 by capillary tube, 30.Bulb 16 throttles valve 14 to regulate temperatures produced inevaporator 15 by the flow of the refrigerant. Passing through evaporator15, the expanded refrigerant absorbs heat returning to the vapor stateaided by circulating fan 32. The refrigerant vapor then returns tocompressor 10 through line 24.

In order to keep the refrigerant in a liquid state in the liquid line,the refrigerant pressure is typically maintained at a high level bykeeping the refrigerant temperature at condenser 11 at a minimum ofapproximately 95° F. This minimum condensing temperature maintainspressure levels in receiver 12 and thus the liquid lines 22 and 23 abovethe flash or boiling point of the refrigerant. At 95° F. condensingtemperature, this pressure for example would be; 125 PSI for refrigerantR12, 185 PSI for refrigerant R22 and 185 PSI for refrigerant R502. Thesetemperature and pressure levels are sufficient to suppress flash gasformation in lines 22 and 23 but the conventional means of maintainingsuch levels by use of high compressor discharge pressures limits systemefficiency.

Various means are used to maintain the temperature and pressure levelsstated above. For example, FIG. 1 shows a fan unit 31 connected tosensor 17 in line 21. Controlled by sensor 17, fan unit 31 is responsiveto condenser temperature or pressure and cycles on and off to regulatecondenser heat dissipation. A pressure responsive bypass valve 18 incondenser output line 21 is also used to maintain pressure levels inreceiver 12. Normally, valve 18 is set to enable a free flow ofrefrigerant from line 21a into line 21b. When the pressure at the outputline of condenser 11 drops below a predetermined minimum, valve 18operates to permit compressed refrigerant vapors from line 20 to flowthrough bypass line 20a into line 21b. The addition to vapor from line20 into line 21b increases the pressure in receiver 12, line 22 and line23, thereby suppressing flash gas.

The foregoing system eliminates flash gas, but is energy inefficient.First, maintaining a 95° F. condenser temperature limits compressorcapacity and increases energy consumption. Although the 95° F.temperature level maintains sufficient pressure to avoid flash gas, theresultant elevated pressure in the system produces a back pressure inthe condenser which increases compressor work load. The operation ofbypass valve 18 also increases back pressure in the condenser. Inaddition, the release of hot, compressed vapor from line 20 into line 21by valve 18 increases the refrigerant specific heat in the receiver. Theadded heat necessitates yet a higher pressure to control flash gasformation and reduces the cooling capacity of the refrigerant, both ofwhich reduce efficiency.

Another approach to suppressing flash gas has been to cool the liquidrefrigerant to a temperature substantially below its boiling point. Asshown in FIG. 1, a subcooler unit 40 has been used in line 22 for thispurpose. However, subcooler units require additional machinery andpower, increasing equipment and operating cost and reducing overalloperating efficiency.

Other methods for controlling the operation of refrigeration systems aredisclosed in U.S. Pat. Nos. 3,742,726 to English, 4,068,494 to Kramer,3,589,140 to Osborne and 3,988,904 to Ross. For example, Ross disclosesthe use of an extra compressor to increase the pressure of gaseousrefrigerant in the system. The high pressure gaseous refrigerant is thenused to force liquid refrigerant through various parts of the system.However, each of these systems is complex and requires extensivepurchases of new equipment to retrofit existing systems. The expensesinvolved in the purchase and operation of these methods usually outweighthe savings in power costs.

A more recent method of controlling the formation of flash gas in theliquid line was disclosed in U.S. Pat. No. 4,599,873 by R. Hyde. Thismethod involves the use of a magnetically coupled centrifugal pumpplaced in the liquid line as seen in FIG. 2. FIG. 2 shows a vapor line114, a condenser 116, a fan unit 118, a liquid line 120, a receiver 122,a pump 124 and 125, a liquid line 126, a heat exchanger 128, a liquidline 129, a valve 130, a line 131, a control 132, an evaporator 134, afan unit 138, and a vapor line 140. The purpose of this method is toimprove system efficiency by allowing system condensing pressures andtemperatures to be reduced as ambient temperatures reduce. Thecentrifugal pump 124 adds pressure to the liquid line 126 at the pointwhere the liquid line exits from the condenser 116 or receiver 122without the use of compressor horsepower. This method of using acentrifugal pump to add pressure reduces the amount of flash gas thatforms in the liquid line, but does not eliminate it altogether.

Furthermore, examination of the centrifugal pump curve in FIG. 3 showsthat as flow increases, the pressure added by the centrifugal pumpdecreases. However, as flow of refrigerant liquid through the liquidline increases the pressure drop in the liquid line increases by thesquare of the velocity. This combination of effects as shown in FIG. 4.causes the centrifugal pump to only reduce the formation of flash gasduring certain low flow conditions, below point A in FIG. 4. Asrefrigerant flow increases at high load conditions and the pressureadded by the centrifugal pump decreases, the formation of flash gasbegins to increase again and system capacity is lost when it is neededmost.

Another deficiency of the previously described centrifugal pumpingmethod is that the centrifugal pump is located within the liquid lineitself. If the centrifugal pump falls to operate properly for anyreason, it becomes an obstruction to flow of refrigerant liquidseriously impairing the operation of the refrigeration system.

The most serious deficiency of the previously describe centrifugalpumping method however, is caused by the state of the refrigerant at theoutlet of the condenser 116 or receiver 122. The liquid refrigerant atthis location in the system is commonly at or very near the saturationpoint. Any vapor that forms at the inlet of the centrifugal pump due toincomplete condensation or slight drop in pressure caused by the pumpsuction or any other reason will cause the centrifugal pump to cavitateor vapor lock and lose prime. This renders the centrifugal pumpineffective until the system is stopped and restarted again, and is verydetrimental to pump life and reliability. Due to the constantly varyingconditions df operation of the refrigeration system this can occur withgreat regularity.

A further development pertaining the fields of mechanical airconditioning and refrigeration relating to system optimization isdisclosed by U.S. Pat. No. 5,150,580 also by R. Hyde. This development,seen in FIG. 2., involves the transfer of some small amount of liquidrefrigerant from the outlet of the centrifugal pump 124 in the liquidline 126 to be injected via conduit 136 into the compressor dischargeline 114 by means of the added pressure of the centrifugal pump 124 inthe liquid line. The purpose of injection this liquid into the dischargeline is to de-superheat the compressor discharge vapors before theyreach the condenser to reduce condenser pressure and thereby reduce thecompressor discharge pressure. This development is said to improvesystem efficiency at high ambient temperatures when air conditioningsystems work the hardest and system pressures are the highest.

Again, however, as system pressures increase and refrigerant flow ratesincrease at higher loads, the increased flow rate of refrigerant causesmore pressure loss through the condenser. However, this same increasedflow rate causes less pressure to be added to the liquid by thecentrifugal pump 124 in the liquid line 126. Thus, less liquid isbypassed via conduit 136 into the compressor discharge line and lesssuperheat is eliminated at the time when more reduction is needed. Andat some point the pressure loss through the condenser is greater thanthe pressure added by the centrifugal pump and the effect is lostentirely.

Obviously, there remains a need to provide a stable pressure increase inthe liquid line 126 to completely eliminate the formation of flash gas,and likewise a stable pressure increase in the liquid injection line 136to completely de-superheat the compressor discharge vapors if theimprovement in system efficiency is to be realized on a constant andreliable basis regardless of system configuration or refrigerant flowrate or vapor content.

There are several major problems associated with adding refrigerantpumps to the liquid line of refrigeration and air conditioning systems:

1. The constantly changing flow rate of the refrigerant.

2. The propensity of refrigerant to boil when it is near saturation asit usually is in these applications.

Centrifugal pumps operate well under a varying flow conditions, but notwhen vapor bubbles form in the liquid as a result of the refrigerantboiling. Then they tend to vapor lock, which prevents them from addingpressure. This makes them unacceptable in refrigerant pumpingapplications since there is a high potential for vapor bubbles to bepresent.

Positive displacement pumps, on the other hand, perform well, even inthe presence of vapor bubbles. This makes them the better choice for usein refrigeration and air conditioning systems. Positive displacement(PD) pumps, however, provide a constant flow rate, so they must bemodified to perform in varying flow rate systems.

The objectives of the present invention are to:

1) Reliably and constantly increase the pressure in the liquid line tosuppress the formation of flash gas without unnecessarily maintaining ahigh system pressure, and without the possibility of obstructing theflow of refrigerant through the liquid line.

2) To reliably and constantly inject the correct amount of liquid intothe compressor discharge line to maximize the heat transfer in thecondenser.

3). To improve the operating efficiency of compression-typerefrigeration and air conditioning systems in a constant, controlled andreliable basis regardless of system configuration or refrigerant flowrate.

4). To maximize the refrigeration capacity of refrigeration and airconditioning systems in a constant, controlled and reliable basisregardless of system configuration or refrigerant flow rate.

5). To economically and constantly suppress the formation of flash gasin refrigeration and air conditioning systems without impairingrefrigeration capacity and efficiency regardless of system configurationor refrigerant flow rate.

6). To provide a way to inexpensively retrofit existing refrigerationsystems to attain the foregoing objects on a reliable and controllablebasis regardless of the system configuration or refrigerant flow rate.

7). To provide a method of automatically reducing the flow rate of thepumping apparatus to match the refrigerant flow rate in largerefrigeration or air conditioning systems that have some unloadingcapability to match the load.

8). Further, the previous objects must be met in a way that will not bedetrimental to the system in the event of failure of the installedpumping mechanism or condenser cooling mechanism.

9). Still further, the above objects must be reliably met regardless ofthe presence of some vapor in the liquid at the inlet of the pumpingarrangement since the liquid is at or near saturation.

10). To assure that pressure added to the refrigerant is accomplishedaccurately and constantly during the widely varying flow conditions ofrefrigerant systems.

11). To virtually eliminate vibration in a positive displacement pumparrangement and avoid cavitation in the liquid line.

12). To allow a positive displacement pump to run substantially unloadedmuch of the time so that pump uses just a fraction of the power it woulduse running at full speed.

13). Moreover, the above objects must be met in a way that can beadjusted to satisfy a majority of the wide range of systemconfigurations found in the field.

This invention provides for the refrigeration or air conditioning systemto be operated in a way which maximizes energy efficiency and suppressesflash gas formation regardless of system configuration or refrigerantflow rate.

The foregoing and other objects, features, and advantages of theinvention will become more readily apparent from the followingdescription of a preferred embodiment, which proceeds with reference tothe figures.

SUMMARY OF THE INVENTION

The invention entails the use of a variable speed drive, positivedisplacement pump magnetically coupled to a drive motor located in aconduit arrangement that is parallel to the liquid line of therefrigeration system as in FIG. 5 This parallel conduit arrangement alsoincludes a pressure regulating valve that will regulate the amount ofpressure added to the liquid line by the parallel pump and pipingarrangement. In addition, a check valve is located in the liquid line tomaintain the pressure differential added to the liquid line. Thisparallel piping arrangement is desirable in order to allow a constant,pre-determined pressure to be added to the liquid line regardless ofvariations in flow rate of the liquid refrigerant. In addition, theparallel piping arrangement allows the system to operate without liquidline obstruction in the event of pump failure.

The present invention involves the use of a controlled variable speeddrive on the pump motor so the flow rate through the pump will moreclosely match the variable system flow rate. This drive may beconfigured for continuously variable speed or a discrete plurality ofspeeds (multiple speed). The term "variable speed drive" in thisdisclosure means either option.

In various embodiments the pump speed can be controlled, continuously ordiscretely by: the amperage draw on a rack of compressors, a signal froma pressure sensor in the liquid line, a signal combined from severalsensors indicating the pressure differential across the pump, a signalfrom a flow sensor in the liquid line at the outlet of the liquidreceiver or condenser, a signal from a pressure or flow sensing devicein a bypass line to vary pump speed to limit the flow of refrigerantthrough the bypass, a signal from a vapor sensing device in the liquidline to vary the pump speed sufficiently to eliminate the vapor, asignal from the refrigeration rack controllers so that pump speed isvaried according to any number of existing inputs, a signal obtained bymeasuring the "conditions " (amount of subcooling) of the refrigerant atthe inlet to the expansion valve, or a signal from a superheat sensor atthe outlet of the evaporator.

Further, a liquid injection line may be added between the outlet of thepump and the compressor discharge line for the purpose ofde-superheating the compressor discharge vapors. The pressure boostprovided by the pump assures a constant flow of liquid refrigerant tothe compressor discharge line. Also, a thermostatic expansion valve isadded at the end of the injection line. Then, a sensing bulb connectedto the thermostatic expansion valve but affixed to the compressordischarge line downstream of the injection point is used to measure thesuperheat and control the operation of the thermostatic expansion valve.In this way the superheat is maximized.

The use of the combination of a positive displacement pump in parallelwith a pressure differential valve is essential to this invention. Theuse of the variable speed drive to control the rate of flow through thepump is also essential in systems where a higher level of control isrequired. The addition of the liquid line and thermostatic expansionline is optional.

The positive displacement pump type of pump is essential for twosignificant reasons, neither one of which can be accomplished with acentrifugal pump.

1. To provide a constant increment of pressure boost over a wide rangeof flow rates.

2. To provide this increment of pressure boost regardless of thepresence of some vapor at the inlet to the pump.

The pressure differential valve is essential in order to limit thepressure boost provided by the pump to a predetermined value.

DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic diagram of a typical refrigeration system, aspreviously described.

FIG. 2 is a schematic diagram of a refrigeration system including theprior art as previously described, including the liquid injection forde-superheating.

FIG. 3 is a diagram of a typical centrifugal pump curve showing pressureadded vs. flow rate.

FIG. 4 is a diagram of pressure loss through a piping system vs. flowrate with the centrifugal pump curve superimposed over it.

FIG. 5 is a schematic diagram of a refrigeration system including anessential precursor to the present invention.

FIG. 6 is a more detailed diagram of the parallel piping arrangementwith positive displacement pump, pressure differential regulating valveand check valves of the precursor to the present invention.

FIG. 7 is a more detailed diagram of the preferred method of addingpressure to the liquid injection line.

FIG. 8 is a diagram of the duplex pumping arrangement used to matchchanging refrigerant flow rate in larger systems with unloadingcapabilities.

FIG. 9 is a blown up depiction of a preferred embodiment of the pump(s)of the present invention.

FIG. 10 shows an earlier development with a fixed speed positivedisplacement pump with a bypass line with pressure differential valve.

FIG. 11 show the use of a variable speed drive controlled by currentbeing supplied to the compressor rack.

FIG. 12 shows a variable speed drive controlled by differential pressuresensors before and after the bypass arrangement.

FIG. 13 shows a variable speed drive controlled by a flow sensor at theoutlet of the receiver or condenser.

FIG. 14 shows a condenser fan deployment controlled by sensors of ampdraw or torque of the variable speed driven pump.

FIG. 15 shows a variable speed drive controlled by a measurement of the"condition", or subcooling, of liquid at the inlet of the expansionvalve.

FIG. 16 shows a variable speed drive controlled by a measure ofsuperheating at the outlet of the evaporator.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

Referring now to FIG. 5, a closed circuit compression-type refrigerationsystem includes a compressor 10, a condenser 11, an optional receiver12, an expansion valve 14 and an evaporator 15 connected in series byconduits defining a closed-loop refrigerant circuit. Refrigerant gas iscompressed by compressor unit 10, and routed through discharge line 20into condenser 11. A fan 31 facilitates heat dissipation from condenser11. Another fan 32 aids evaporation of the liquid refrigerant inevaporator 15. The compressor 10 receives warm refrigerant vapor atpressure P1 and compresses and raises its pressure to a higher pressureP2. The condenser cools the compressed refrigerant gases and condensesthe gases to a liquid at a reduced pressure P3. From condenser 11, theliquefied refrigerant flows through line 21 into receiver 12 in caseswhere there is currently a receiver in the system. If there is noreceiver in the system the condensed refrigerant flows directly into theliquid line 22. Receiver 12 in turn discharges liquid refrigerant intoliquid line 22.

FIG. 6 shows a positive displacement pump 41, driven by electric motor42 magnetically coupled to the pump head is positioned in conduitarrangement 60 parallel to the liquid line 22 at the outlet of thereceiver or condenser to pressurize the liquid refrigerant in the lineto an increased pressure P4. This parallel piping arrangement 60 alsoincludes the pressure differential regulating valve 45 and a check valve46 arranged as shown in FIG. 6 to provide for a constant added pressure(P4-P3) regardless of refrigerant flow rate or vapor content. A checkvalve 47 is added to the liquid line 22 to maintain the pressuredifferential between line 22 and line 23 (see FIG. 7). An adjustablepressure regulating valve 45 can also be used to more accurately matchthe pressure differential be required or to facilitate changes that maybe needed in the pressure differential added. The pressure differentialof the regulating valve 45 (FIG. 6) determines the amount of pressurethat is added to the system. Different amounts of pressure can be addedto the liquid line 22 as necessary for each different systemconfiguration by using different pressure differential valves or byadjusting the valve to a specific pressure as needed. As the flow rateof the system varies in conduit 22, more or less refrigerant flowsthrough parallel conduit 22a (FIG. 6) and pressure regulating valve 45so the refrigerant flow into and out of the parallel piping arrangement60 always matches the flow rate through conduit 22 and 23 and thepressure differential (P4-P3) remains constant.

From parallel piping arrangement 60, the liquid refrigerant flows intothe liquid line 23 (FIG. 7). Some of the liquid refrigerant flowsthrough conduit 25 and thermostatic expansion valve 81 into thecompressor discharge line to de-superheat the compressor dischargevapor. The thermostatic expansion valve is controlled by bulb 48 whichsenses the temperature of the superheated vapor.

The remainder of the liquid refrigerant from the parallel pipingarrangement 60 (FIG. 5) flows through the line and through an optionalcounter-current heat exchanger (not shown) to thermostatic expansionvalve 14. Thermostatic expansion valve 14 expands the liquid refrigerantinto evaporator 15 and reduces the refrigerant pressure to near P1.Refrigerant flow through valve 14 is controlled by temperature sensingbulb 16 positioned in line 24 at the output of evaporator 15. Acapillary tube 30 connects sensing bulb 16 to valve 14 to control therate of refrigerant flow through valve 14 to match the load at theevaporator 15. The expanded refrigerant passes through evaporator 15which, aided by fan 32, absorbs heat from the area being cooled. Theexpanded, warmed vapor is returned at pressure P1 through line 24 tocompressor 10, and the cycle is repeated.

Pump 41 and pressure regulating piping arrangement 60 is preferablylocated as close to receiver 12 or the outlet of condenser 11 aspossible, and may be easily installed in existing systems withoutextensive purchases of new equipment. Pump 41 must be of sufficientcapacity to increase liquid refrigerant pressure P3 by whatever pressureis necessary to eliminate the formation of flash gas in the liquid line23 (FIG. 7). The pump must also be capable of adding a constant pressureto the liquid line regardless of the presence of some vapor in theincoming liquid refrigerant in line 22. A positive displacement pump andpressure regulating valve located in a parallel piping arrangement 60most effectively, economically and reliably provides this capability.

Pump 41 must also be capable of adding a constant pressure to the liquidline under conditions of variable refrigerant discharge rates from valve14, including conditions in which valve 14 is closed.

In systems where the refrigerant flow rate varies significantly, thepumping arrangement must be able to vary its flow rate by a similaramount. In these cases, a duplex pumping arrangement, FIG. 8, may beused. The duplex pumping arrangement consists of two pumps piped inparallel each with either a single speed, two speed or variable speedmotor and a control mechanism capable of adjusting the speed of one orboth pumps to match the flow rate of the refrigerant in therefrigeration circuit. This duplex pumping arrangement is typically usedin systems that have multiple compressors or compressors with thecapability of unloading to significantly reduce the refrigerant flowrate. The duplex pumping arrangement controls tie into the systemcontrols to adjust the pump or pumps speed to match the compressorloading thereby matching the refrigerant flow rate.

There are several possible modifications to the installation of thepositive displacement pumps which allow us to most efficiently takeadvantage of their superior performance with saturated liquids.

1. A bypass with a pressure differential check valve has been added toinsure that a predetermined pressure differential exists across thepump, and that there is a path for excess flow, and

2. A variable speed drive has been installed on the positivedisplacement pump motor so the flow rate through the pump will moreclosely match the system flow rate.

In some large systems, refrigeration system racks are comprised ofseveral compressors manifolded together and sized to handle the maximumload of the system. The compressors cycle off and on individually tomatch the varying system load. At any given time the system load andresulting refrigerant flow rate may be at its maximum or none, oranywhere in between. Since, the positive displacement pump flow ratemust also be designed for maximum load of the system, under all but themost loaded conditions, the positive displacement pump will beoversized.

The theory behind the application of a variable speed drive to apositive displacement pump is that by controlling the speed of the motordriving the positive displacement pump, the flow rate will vary directlywith it. This concept has been tested on a large supermarketrefrigeration system. In order to vary the positive displacement pumpflow rate with the flow rate of the system, an amperage sensor wasplaced onto the main electrical feed to the rack between the powersource and the compressor rack. The system, and the refrigerant flowrate increases and decreases with the load in a similar fashion. Theamperage sensor sensed the current draw of the system and produced avariable signal output based on this current draw. This variable signaloutput was sent to the signal input of the variable speed drive therebyvarying the speed of the drive and pump motor based on the current useof the rack (FIG. 11).

Using this method, we found that the advantages of using the positivedisplacement pump could be fully realized.

1. The pressure added to the refrigerant was accurate and constantduring the widely varying flow conditions of the refrigeration system.

2. Much less refrigerant was bypassed through the overflow valve. Thisvirtually eliminated vibration in the positive displacement pump andcavitation in the liquid line.

3. Since the positive displacement pump ran substantially unloaded muchof the time, the pump itself used just a fraction of the power that itwas using when it was running at full speed.

It should be noted that a number of alternative input devices could beused to control the speed of the motor, and therefore the refrigerationflow.

1. A pressure sensing device could be used to provide a constant liquidline pressure.

2. Several sensing devices could be used to provide a constant pressuredifferential across the pump (FIG. 12).

3. A pressure or flow sensing device in the bypass line could be used tovary the speed of the pump to limit the flow of refrigerant through thebypass.

4. A vapor sensing device in the liquid line could be used to vary thespeed of the pump sufficiently to eliminate the vapor.

5. The variable speed drive could be tied into refrigeration rackcontrollers which would vary the speed according to any number ofexisting inputs.

In addition to these, there are a great many other devices and means ofcontrolling the speed of the pump and associated pressure boosts andflow rates.

Operation

Referring to FIG. 5, compressor 10 compresses the refrigerant vaporwhich then passes through discharge line 20 to condenser 11. In thecondenser 11, at pressure P2, heat is removed and the vapor is liquefiedby use of ambient air or water flow across the heat exchanger. Atcondenser 11, temperature and pressure levels are allowed to fluctuatewith ambient air temperatures in an air-cooled system, or with watertemperatures in a water-cooled system to a minimum condensingpressure/temperature that has previously been set at about 95° F. Thispreviously set minimum condensing temperature has been necessary toprevent the formation of flash gas in the liquid line 22. The previouslyset minimum was maintained by reducing air or water flow across the heatexchanger of condenser II to reduce heat transfer from the condenser.Further decreasing the condensing temperatures increase systemefficiency in two ways: 1) The lower pressure differential of thecompressor 10 increases the compressor volumetric efficiency accordingto the formula V_(e) =1+C-C*(V₁ /V₂) where V_(e) is volumetricefficiency, C is the clearance ratio of the compressor, V₁ is thespecific volume of the refrigerant vapor at the beginning ofcompression, V₂ is the specific volume of the refrigerant vapor at theend of compression, and 2) The lower liquid refrigerant temperature atthe outlet of the condenser results in a greater cooling effect in theevaporator.

The negative effect of reducing condensing temperatures below thispreviously set minimum has been the formation of flash gas in the liquidline 23 (FG. 7), which when passed through expansion valve 14 reducedthe net refrigeration effect of the evaporator 15. The net result was areduction of energy consumption per unit time by the compressor, but asimultaneous reduction capacity of the system causing an increase incompressor run time resulting in no net energy savings.

When the refrigeration or air conditioning system is modified with thepresent invention as in FIG. 5, the minimum condensing temperature andpressure can be reduced significantly without the loss of capacitymentioned above due to the pressure added to the liquid line by the pump41 and parallel piping arrangement. As the ambient air temperature orwater temperature used to cool the condenser becomes lower, theefficiency of the compressor improves, and the capacity of theevaporator increases, since no flash gas has been allowed to form in theliquid line. This is most beneficial with refrigeration systems thatoperate year around and can take advantage of the cooler ambienttemperatures.

As ambient air temperature or cooling water temperature increases thecondensing temperature and pressure of the refrigeration or airconditioning system also increases and efficiency is reduced. In orderto improve efficiency at these higher ambient conditions when airconditioning and refrigeration systems are at or near maximum capacity,liquid refrigerant is bypassed from the liquid line 23 (FIG. 7) into thecompressor discharge line 20. Since there is some amount of pressurelost as the refrigerant passes through the condenser 11, makingcondenser exit pressure P3 lower than entrance pressure P2, a pressureboost is needed to insure flow of liquid from the liquid line 23 intothe discharge line 20. Pump 41 provides this pressure boost.

An alternative method is to use a separate positive displacement pump43, driven by a variable speed drive, controlled by the temperaturedifferential between the superheated compressor discharge vaportemperature 12 and the condensing temperature T3. As the temperaturedifferential becomes greater, the variable speed drive would cause thepositive displacement pump to pump more liquid into the discharge line20 to decrease the superheat. When the superheat temperature and thecondensing temperature were the same, the refrigerant vapor entering thecondenser would be at the saturation point and the speed of the positivedisplacement pump would stabilize to a pre-set speed to maintain thecondition.

This method of superheat suppression insures that the refrigerant vaporis entering the condenser at saturation resulting in the optimumconditions for heat transfer thereby optimizing the efficiency of thecondenser. This portion of the invention is most beneficial at higherambient temperature.

Referring to FIG. 9, the pump(s) of the present invention consists of anouter driving magnet 200, a stationary cup 201, and an O-ring seal 202.The pump further includes an inner driven magnet 203, a rotor assembly204 and vanes 205. The pump further includes an O-ring seal 206 andbrass head 207.

Taken together, both parts of the invention improve system performanceand efficiency over the full range of operating conditions andtemperatures.

The use of magnetically-coupled rotary-vane pumps as positivedisplacement pumps for pumping refrigerants has been found to bestartlingly effective and they have been found to exhibit a surprisinglylong life. Once the vanes are worn to the extent that they are properlyseated and sealed, subsequent wear is almost negligible. This discoveryhas resulted in very effective use of these magnetically-coupledrotary-vane pumps as positive displacement pumps for pumpingrefrigerants in non-compressor-type refrigeration cycles. Thisapplication is particularly effective when a compressor-typerefrigeration cycle (preferably with the help of the present invention)is used to store refrigeration, for example, in the form of ice, duringlow energy cost periods and then the compressor is turned off duringpeak energy cost periods. During the peak period, themagnetically-coupled rotary-vane pump of the present invention (ideallythe same pump used to increase the efficiency of the compressor cycle)is used to circulate the same refrigerant through the ice, through thesame conduits, and through the same cooling coils (evaporator), to coolthe conditioned space during peak energy cost periods.

Another aspect of the present invention is the use of staring torquecontrol means for the positive displacement pump. Typically, when apositive displacement pump is placed into the liquid line of an airconditioning or refrigeration system, the electric motor driving it isenergized when the compressor is energized. This creates two problemswhen the pump head is full of refrigerant upon start-up, as it isnormally the case. First, excessive torque is required to bring the pumphead up to speed while it is adding pressure to the liquid. Second, therapid acceleration of the pump rotor will cause temporary, butsignificant, cavitation that may damage the pump.

The solution to both problems is to ramp the motor and pump up tooperating speed slowly. This can be accomplished by using a devicecalled a "soft starter". This device will bring the motor up to fullspeed over a period of 1 or more seconds, depending on its design.

Upon normal start-up, a standard electric motor will go from 0 R.P.M. toits full speed of 3450 R.P.M. in less than 1 second. This causesexcessive torque requirements and cavitation when such a motor iscoupled to a positive displacement pump that is full of a liquid nearsaturation. If the acceleration rate of the motor and pump head isslowed down so that it comes up to speed in preferably between 2 and 8seconds, for example, the excessive torque and cavitation problems areavoided.

The variation in start-up acceleration can be accomplished by severalmeans: 1. using an induction coil in series with the electric motor, or2. redesigning the motor windings to give less start-up torque,therefore slower starting speed, or 3. installing a separate "softstart" electronic component to a standard motor that varies the voltageto the motor.

The type of pump is important, contrary to what the prior art teaches,and it must be a positive displacement type contrary to what isdisclosed in prior art systems.

In order to insure stable and therefore optimal system operation, thepressure valve must be a pressure differential valve not a pressurelimiting valve as shown in prior art. The purpose of the added pressureis only to overcome the pressure loss in the liquid line to prevent theformation of vapor in the liquid line. The pressure differential valve,set at a constant, predetermined pressure differential accomplishes thiswithout the use of excess pumping energy. The pressure limiting valve ina prior method, limits the reduction of pressure in the liquid line.This method holds excess pressure in the liquid line during periods oflow condensing pressure, but does nothing to prevent vapor formation inthe liquid line during periods of higher condensing pressure. Thepurpose of the prior pressure limiting valve method is to maintain ahigh pressure differential across the metering device at the inlet tothe evaporator. The purpose of the pressure differential valve of thepresent, improved method is to maintain optimum metering device capacityby constantly adding the predetermined pressure necessary to prevent theformation of vapor in the liquid line during all periods of operation.

To further optimize system performance a variable speed drive is used tovary the flow rate of the positive displacement pump while maintaining aconstant pressure differential.

Three factors effect the capacity of the system refrigerant meteringdevice and therefore the capacity of the system; 1) Quality of liquidrefrigerant at the inlet to the metering device. ie.: If any vapor ispresent in the liquid refrigerant entering the metering device, thesystem capacity is reduced by the percent of vapor present, 2)Temperature of the refrigerant at the inlet to the metering device. ie:The lower the temperature of the refrigerant, the higher the capacity ofthe metering device, and 3) The pressure differential across themetering device. ie: The lower the pressure differential across themetering device, the lower its capacity.

In order to optimize system performance, all three factors must besimultaneously and constantly controlled.

1. The use of a non-centrifugal type of pump, preferably a positivedisplacement type of pump, is necessary in the scope of the currentinvention to provide a constant, predetermined increment of pressure tothe liquid refrigerant in the liquid line 22. In refrigeration or airconditioning systems, the flow rate of the refrigerant within the systempiping varies continuously as the cooling load on the system varies. Inorder to provide a constant increment of pressure regardless of thesystem flow rate, a positive displacement pump must be used inconjunction with a bypass line (22B) with a pressure differential valveas shown in FIG. 10. The positive displacement pump provides a fixedflow rate that is higher than the flow rate of the system. The bypassline provides a path for the difference in flow between the constantlyvarying system flow rate and the fixed pump flow rate. In that way, theflow rate of refrigerant into and out of the bypass arrangement isalways exactly matching the flow rate of the system, while the flow rateof the refrigerant through the positive displacement pump and thepressure added by the pump remain constant.

2. A pressure differential valve is used in the bypass line 22B toprovide the constant increment of pressure necessary to satisfy therefrigerant metering device.

The temperature and pressure of the refrigerant in the condenser varytogether as the refrigerant is condensing from a vapor to a liquid. Asthe temperature of the condensing medium is reduced, the temperature andpressure of the refrigerant being condensed to a liquid can be reduced.The result is, as the condensed refrigerant liquid temperature isreduced, its pressure is also reduced. Since the capacity of themetering device increases with a reduction in liquid temperature anddecreases with a reduction in liquid pressure, the net capacity of therefrigerant metering device will remain relatively constant as long asthe temperature and pressure differential are reduced together, andthere is no vapor present in the liquid line or at the inlet to themetering device.

The pressure differential valve allows this reduction in temperature andpressure to occur while the pump adds the minimum constant increment ofpressure necessary to prevent vapor form forming in the liquid line. Theaddition of the lowest constant increment of pressure necessary insteadof adding excess pressure, up to the limit of the pressure regulatingvalve, allows for the optimal system operation to be maintained withoutthe use of excess energy that would be required to add the excesspressure up to the limit of the pressure regulating valve.

FIG. 10

Shows the processor to the present invention with constant speed driveto provide steady flow rate and pressure increment with the differencein flow rates between pump flow rate and system flow rate being bypassedthrough line 22B. This method allows for a constant, predeterminedincrement of pressure to be added while the flow rate through the bypassarrangement exactly matches the varying system flow rate through line22. This method would be used in refrigeration and air conditioningsystems where the variation in refrigerant flow rate is not great andthe compressor cycles on and off to match the system load. In thismethod, the pump is energized whenever the compressor is energized.

FIG. 11

In many larger refrigeration and air conditioning systems, the system isdesigned to have the capacity necessary to satisfy the maximum loadrequired, but the actual load on the system is significantly lower thanthis maximum during a majority of its operating hours. By the sametoken, the refrigerant pumping system must be sized for the maximumrefrigerant flow rate, but the actual refrigerant flow rate issignificantly lower than this maximum during most of its operatinghours.

In these larger refrigeration and air conditioning systems, therefrigerant flow rate is varied while the compressor or compressorsremain energized. This is done by either using multiple compressors thatcycle on and off as needed to match the load on the system, or by usinga single compressor with several cylinders that are activated ordeactivated as needed to match the load on the system. In systems suchas these, a variable speed drive is used to drive the positivedisplacement pump. The speed of the pump motor, and therefore the flowrate of the pump can be regulated by some signal from the system so theflow rate provided by the pump more closely matches the flow rate of thesystem.

The purpose of this invention is to optimize the efficiency of theoperation of the standard refrigeration cycle. Likewise, the purpose ofthe variable speed drive is to optimize the efficiency of the operationof the refrigerant pump. Optimal pump operation is that which consumesthe least amount of energy necessary to add the predetermined incrementof pressure to the liquid line. The point of "least amount of energynecessary" occurs just as the pressure differential check valve is inthe bypass line begins to open. Just before this point, the pressureadded by the pump is not as high as the predetermined set point of thepressure regulating check valve. Just after this point, liquid begins toflow through the bypass and is recirculated by the pump requiring morework to be done by the pump than is necessary. Ideally then, the speedof the pump should be varied with the refrigerant flow rate to justmatch the flow required to start to open the pressure differential valvein the bypass line, and no more.

The preferred method of varying the flow rate of the positivedisplacement pump to more closely match the system flow rate in order tominimize the excess flow through the bypass line 22B in systems wherethe compressor or compressors operate continuously, and some means ofcompressor unloading occurs, is shown in FIG. 11. The flow rate providedby a positive displacement pump varies directly with the rotationalspeed of the pumping mechanism. Therefore, if the speed of the motordriving the pump is varied, the flow rate provided by the positivedisplacement pump can be varied at a predetermined rate.

In the preferred method shown in FIG. 11, an electrical current sensor(71) is attached to the wires that supply the refrigeration or airconditioning system compressor or compressors (10). As the load on thesystem compressors varies, the current required by the compressorsvaries. This variation in current is measured and a variable outputsignal that varies as the system current use changes is provided by thecurrent sensor. This variable output signal is fed through wire 80 tothe controls of a variable speed drive (72) attached to the pump motor.As the current required by the compressors varies, the signal outputfrom the sensor changes the speed of the motor driving the pump therebycausing the flow rate of the pump to vary with the load on thecompressors.

For example, the maximum current required by a refrigeration system atfull load is 100 amps, and varies with load down to 0 amps when thesystem is off. A current sensor that generates a 4 to 20 milliampcontrol signal is attached to the electrical wires that energize therefrigeration system. If the system is operating at full load and isdrawing 100 amps, the amperage sensor generates a 20 milliamp signaloutput. If the system is off and is drawing 0 amps, the amperage sensorgenerates a 4 milliamp output signal. This signal is fed by means of acontrol wire to the control input of a variable speed drive controllerthat controls the speed of the pump. If the variable speed drive controlis fed 20 milliamps, the pump operates at full speed. If the variablespeed drive control is fed 4 milliamps, the pump will not operate. Thespeed of the pump then varies linearly with the 4 to 20 milliamp signalto match the load on the compressors and therefore the refrigerant flowrate.

FIG. 12

Another method of varying the flow rate of the pump to more closelymatch the flow of refrigerant in the system is shown in FIG. 12. Twopressure sensors, 73 and 74 are attached the liquid line. One of thesesensors measures the pressure in the liquid line before the bypassarrangement, pressure P3 and the other measures the pressure in theliquid line just after the bypass arrangement, pressure P4. These twopressure sensors are connected to the pressure regulator 75. Thepressure regulator is set to control the pressure differential to apredetermined differential, PD1, as required by the pressure loss in theliquid line between the condenser or receiver and the refrigerantmetering device. The pressure controller generates an output controlsignal that varies linearly as the difference between the presetdifferential PD1 and the measured pressure differential P4-P3 varies.This variable output signal is input into the controls on a variablespeed drive 72. As the pressure differential between the two sensorsPD4-PD3 increases above the preset amount PD1, the pressure controllerreduces the signal fed to the variable speed drive, and the variablespeed drive reduces the speed of the pump until the preset pressuredifferential PD1 is reached. If the measured pressure differentialPD4-PD3 is less than the preset pressure differential PD1 the pressurecontroller increases the signal fed to the variable speed drive, therebyincreasing the speed of the pump.

FIG. 13

Another method of varying the flow rate of the pump to more closelymatch the flow of refrigerant in the system is shown in FIG. 13. A flowsensor F1 is placed in the liquid line of the refrigeration system 22 atthe outlet of the liquid receiver or condenser. The sensor measures theflow of refrigerant and generates a varying output signal that varieslinearly with the variation in refrigerant flow rate. This varyingcontrol signal is input to a variable speed drive (72) which drives thepump motor. As the refrigerant flow varies, the control signal from theflow sensor varies and changes the speed of the variable speed drive.This in turn varies the speed at which the pump is operated varying theflow of refrigerant through the pump.

FIG. 14

In order to take advantage of the energy savings possible when employingthe current invention, the refrigeration or air conditioning systemcondensing pressure/temperature is allowed to float lower than thenormal factory preset levels. There is a potential for system capacityloss if the pump fails to add pressure to the system when the condensingpressure/temperature is lower than normal. In order to prevent this fromoccurring when the pump fails to add pressure, the system condensingpressure/temperature control can be raised to its original setting. Thiscan be done with the pump motor variable speed drive mechanism (72).When this mechanism senses a significant reduction of pump motor ampdraw or pump torque, it will sent an output signal to the condenser fancontrols that will switch them back to their original setting.

System condensing pressure in air cooled systems is controlled bycycling the condenser fans on and off to maintain whatever minimum isrequired. In order to lower the condensing pressure/temperature, thefans are turned on. In order to maintain or raise the condensingpressure/temperature, the fans are turned off.

FIG. 15

Another method of varying the flow rate of the pump is to measure thecondition of the refrigerant at the inlet to the TXV 14. Since thepurpose of the present invention is to add pressure to the liquid lineto properly feed liquid refrigerant at the proper condition to the TXV,that condition at the inlet of the TXV can be monitored and an outputsignal sent back to the pump to vary its speed.

The condition (amount of subcooling) of the refrigerant at the inlet tothe TXV can be determined by monitoring its pressure and temperature asshown in FIG. 15. A pressure sensor P and the temperature sensor T areattached to the liquid line 22 very near the TXV 14. These sensorsoutput either a mechanical or electrical signal to signal analyzer 73that in turn sends an output signal to the variable speed drive 72 ofthe pump motor based on a preset minimum pressure and temperaturecondition. As the amount of subcooling sensed at the inlet to the TXVreduces, the speed of the VSD would increase thereby increasing thepressure in the liquid line and increasing the subcooling.

FIG. 16

Still another method of varying the flow rate of the pump to match thesystem flow rate is by using a superheat sensor similar to the existingTXV sensing bulb 16. The increase or decrease in pressure in the sensingbulb capillary tube resulting from the increase or decrease in superheatat the outlet of the evaporator acts to move a diaphragm in the controlmechanism 73. This movement is translated into an output signal that isin turn fed into the variable speed drive 72 for the pump motor. Thehigher the superheat sensed by the bulb 16B, the faster the pump motoris turned. This will add more pressure to the liquid line which willfeed more liquid into the TXV and the evaporator which will in turnlower the superheat at the sensing bulb 16B. The motor speed will thenmodulate continuously to hold the superheat to some preset conditionsimilar to the way TXV sensing bulb 16 modulates the TXV.

In addition, there can be any number of different sensor inputs to thesignal analyzer and/or controller 73 based on different system variablesto control the pump speed for a particular application.

Having described and illustrated the principles of the invention in apreferred embodiment thereof, is should be apparent that the inventioncan be modified slightly in arrangement and detail without departingfrom such principles. In that regard, this patent covers allmodifications and variations falling within the spirit and scope of thefollowing claims:

We claim:
 1. Any refrigeration, air conditioning or process coolingsystem using a reciprocating screw, scroll, centrifugal or other similartype of compressor and any type of refrigerant,the improvement includinga positive-displacement pump used in a parallel piping arrangement whicharrangement is parallel to a conventional liquid conduit between acondenser and an expansion valve, and parallel with a check valve, avariable speed drive, driving said positive displacement pump, and adrive controller connected to and controlling said variable speed driveand having as input a signal from a sensor of a variable proportional torefrigerant flow in the system or a related variable, whereby the speedof the positive displacement pump is adjusted to the minimum speednecessary to add a predetermined increment of pressure to the liquidconduit or to eliminate flash gas.
 2. A system as recited in claim 1,wherein the system includesa control system which sets the minimumcondensing temperature setting of refrigerant exiting the condenser to alower-than-conventional value when the pump is functioning properly andreverts the air conditioning or refrigeration system back to the higherminimum condensing temperature setting in case of failure of the pump.3. A system as recited in claim 1 further characterized by the provisionof:a compressor rack having an electrical power source, and a sensor ofamperage draw by the compressor rack producing a signal proportional tosaid amperage draw and communicating with said drive controller tocontrol said pump speed.
 4. A system as recited in claim 1 furthercharacterized by the provision of:a pressure sensor in said liquidconduit producing a signal proportional to said pressure andcommunicating with said drive controller to control said pump speed. 5.A system as recited in claim 1 further characterized by the provisionof:a pair of pressure sensors at, respectively, the input and output ofthe pump assembly producing a combined signal proportional the pressuredifferential across the pump and communicating with said drivecontroller to control said pump speed.
 6. A system as recited in claim 1further characterized by the provision of:a flow sensor in the liquidconduit at the outlet of the liquid receiver or condenser producing asignal proportional to the liquid flow rate and communicating with saiddrive controller to control said pump speed.
 7. A system as recited inclaim 1 further characterized by the provision of:a vapor sensor in theliquid conduit communicating with said drive controller to control saidpump speed sufficiently to eliminate the vapor.
 8. A system as recitedin claim 1 further characterized by the provision of:a compressor rackhaving an electrical power source and a rack controller, said rackcontroller communicating with said drive controller to control said pumpspeed according to the same inputs received by said rack controller. 9.A system as recited in claim 1 further characterized by the provisionof:a sensor of the amount of subcooling of the refrigerant at the inletto the expansion valve and communicating with said drive controller tocontrol said pump speed.
 10. A system as recited in claim 1 furthercharacterized by the provision of:a superheat sensor at the outlet ofthe evaporator providing a signal proportional to the degree ofsuperheat and communicating with said drive controller to control saidpump speed.
 11. A system as recited in claim 1, wherein the systemincludesa liquid injection line between the output of the pump and theoutput of a compressor, used for de-superheating the compressordischarge vapor, and a thermostatic expansion valve and sensing bulb tocontrol the flow of liquid refrigerant through the injection line.
 12. Avapor-compression heat transfer system having fluid refrigerant, acompressor, a condenser, an expansion valve, an evaporator, arefrigerant conduit between the condenser and the expansion valve, and arefrigerant pump in the conduit adapted to increase the pressure of therefrigerant between the condenser and the expansion valve,theimprovement comprising(a) the fact that the said pump is a positivedisplacement pump, and (b) a bypass conduit is provided in parallelaround the pump, said bypass conduit including a check valve adapted tostop flow of refrigerant through the said bypass conduit from theexpansion valve to the condenser, but to allow flow of refrigerantthrough the said bypass conduit from the condenser to the expansionvalve, (c) said pump and bypass conduit being adapted to increase thesaid pressure of the refrigerant-sufficiently to avoid the formation ofrefrigerant flash gas in said conduit between the pump and the expansionvalve, while still allowing flow of refrigerant from the condenser tothe expansion valve if the pump fails to operate, (d) a variable speeddrive, driving said positive displacement pump, and (e) a drivecontroller connected to and controlling said variable speed drive andhaving as input a signal from a sensor of a variable proportional torefrigerant flow in the system or a related variable, whereby the speedof the positive displacement pump is adjusted to the minimum speednecessary to add a predetermined increment of pressure to the liquidconduit or to eliminate flash gas.
 13. A vapor-compression heat transfersystem as recited in claim 12, wherein a liquid injector conduit isprovided between an output side of the pump to an output side of thecompressor, and adapted to deliver pressurized liquid refrigerantde-superheat the refrigerant when it exits the compressor.
 14. Avapor-compression heat transfer system as recited in claim 13, whereinthe liquid injector conduit includes a thermostatic expansion valve andbulb sensor to monitor the temperature of the gas exiting the compressorso as to minimize the superheat in the refrigerant.
 15. Avapor-compression heat transfer system as recited in claim 13, wherein acontrol system is provided to cause reduction in the minimum condensingtemperature at the outlet of the condenser when the pump is effectivelyreducing flash gas, but the control system is adapted to raise theminimum condensing temperature to a point which reduces flash gas, ifthe pump fails to operate.